Flame Imaging Studies of Cycle-by-Cycle

Combustion Variation in a SI Four-Stroke Engine

ABSTRACT

Sets of sequential-cycle instantaneous flame images are used to study
cycle-by-cycle variation of lean combustion in a spark-ignition four stroke
optical engine. Stereo gated image-intensified NTSC video cameras record flame
radiation to show three-dimensional structure, while flame development is
measured in each cycle by superimposing early and later flame images.

A variety of physical causes for cyclic variation are identified from the
images. Correlations of flame geometry with mass burned fraction show that a
larger initial flame kernel size results in a faster early burn in the cycle,
and that flames that are flatter on a large scale cause lower peak burn rates
than round flames. The early flame kernel is shown to vary greatly in size,
shape, and location. This kernel has a major effect on combustion by setting the
basic flame shape in mid-cycle. Large scale unstable flows appear to cause major
cyclic variation in flame shape and combustion, an effect seen at 500 rpm, but
which has disappeared at 1000 rpm.

SAE Paper 892086 (1989)

INTRODUCTION

Cyclic variation of automobile engine combustion is a fundamental
characteristic of the powerplant that is the primary means of transportation in
this country. The extremes of this variation reduce drivability and gas mileage,
and can be responsible for significant air pollutant emissions from the engine.

There have been many studies of cyclic variation from various perspectives. A
review has been done by Young [1], and some of the more recent studies may serve
as a guide to ongoing work ([2], [3], [4] and [5]).

Various heuristic procedures have been used to design engines for
"improved" combustion. The intention is to achieve combustion that has
optimum air/fuel characteristics and optimum phasing with minimum duration and
cyclic dispersion over the full range of engine operating conditions. For many
parameters some of these requirements are conflicting. For example, high swirl
engines minimize cyclic variability but have reduced charging characteristics.

At this time, studies of turbulent flows and turbulent combustion are done
primarily in a statistical manner, as are studies of cyclic variation in
automotive engines. This may be appropriate for describing the overall system
behavior of an engine, but such statistics depend on engine configuration and
operating condition. Much of past work has been to correlate engine parameters
with output statistics [1,7], with decidedly mixed success. As in the case of
any probabilistic set of events, each individual example has its details that
contribute to a unique event. In congregating engine statistics, various
contending effects are usually combined to give data that often have little
correlation with any but gross physical effects.

A major purpose of the present work is to begin to identify independent
physical causes of cyclic variation by examining the development of the flame in
each cycle. A complementary task is to identify
effects that result in misleading statistics. In this work, light-amplified
flame images of many consecutive operating cycles from a realistic optical
engine [9] provide one means among many [10] of exploring cycle
by-cycle variation in some different engine configurations. Previous work [8]
at retarded spark timing provides a foundation for the present study.

EQUIPMENT AND PROCEDURE

Since cyclic variation in turbulent combustion is dependent on so many
factors, it is necessary to specify as much of any experiment system as
possible. Subtle details often become critical under some operating conditions.
In the present work lean combustion is studied, where flame speeds more closely
approximate flow velocities, and cyclic variation is usually greatly enhanced.

ENGINE - The flame image data of this work were taken from a single-cylinder
visualization engine described in detail elsewhere [8,9]. The major
characteristics of this engine are representative of a modern-four-stroke
spark-ignition internal combustion engine; global parameters for the engine are
given in Table 1. The engine has a sapphire liner with a flat quartz piston-top
window in an extended piston. Premixed propane and air is drawn through the
engine for a firing run from an intake tank of 50 cylinder-volumes.

Shown in Figure la is the pent-roof head geometry and the transparent liner
that extends to the bottom of the head. Figure lb shows the valve, spark plug,
and pressure transducer location in the head. Although the head is designed for
two spark plug operation, the secondary plug has been replaced with a pressure
transducer for this work. The peripheral spark plug is oriented vertically
(parallel to the cylinder axis) and is electrically driven by a standard HEI
ignition system. The piston window diameter is 69.8 mm at its smallest section,
compared with the 92 mm, cylinder bore. Combustion chamber crevices consist of a
0.6 mm gap 5.0 mm deep around the piston, and a 3 mm deep, 0.1 mm wide gap at
the bore radius in the head.

Two basic engine configurations were used. Although both valves operate
normally, the secondary port can be throttled at the entrance to the head. This
throttle was closed for data referred to as "l-port", resulting in a
mean swirl ratio of 1. Data referred to as "2-port" differs from this
only with respect to the opening of this throttle plate, giving very little
swirl. In all cases the manifold absolute pressure (MAP) was adjusted to 0.75
atm.

FLAME IMAGING DIAGNOSTIC - Flame behavior is studied by analyzing short
exposure stereo images of flame luminosity taken by two image-intensified CCD
video cameras. The flame imaging system is described in detail elsewhere [8,9].
The cameras have a linear response to visible light and a luminous gain variable
up to 20,000. The basic format of a digitized flame image is an array of 512 x
512 pixels, with an 8 bit intensity resolution. Spatial resolution of the images
is on the order of a few pixels, limited by the image intensifier resolution,
decreasing with increasing gain. Flame images are analyzed by a commercial image
processing package on a high-resolution (1024 x 1280), 256 color IBM PC/AT
monitor. The images are correlated with pressure data that has been
post-processed to give mass burned progress using techniques similar to those
described in Rassweiler and Withrow [11].

FLAME IMAGE DATA EXPLANATION - The flame image data presented in this work
gives information about both the spatial and temporal structure of the flame. A
sample of this data is shown in Figure 2.

Spatial variation of the intensity of the flame radiation integrated along a
line of sight is shown for the bottom view through the piston top. As for most
of the figures, a line drawing derived from a reference image shows the
background engine geometry, superimposed on the flame image. The line drawing of
Figure 2 reflects the geometry shown in Figure 1b, including the valves and
spark plug with the electrode at the lower end of the ground strap. The
protruding shape between the intake valves is the bottom flat of the head
protruding into the window view. These line drawings are added with image
processing software for each sequence of cycles without any adjustment of their
relative location.

Flame development is recorded during a single cycle at normal engine speeds
by making two sequential exposures at different times on one frame, necessitated
by the 30 frames/sec standard video framing rates. Single image exposures as
short as a few microseconds are obtained by gating the image intensifier. For
the case of peripheral ignition described in this work, the early flame (bright
patch close to the spark plug in Figure 2) is spatially separated from the flame
front at TDC which has moved to a different part of the cylinder. The images in
Figure 2 were taken at 20° BTDC and at TDC, each approximately 1°
in duration. Since the flame front radiation is much greater than the
comparatively uniform and weak radiation from the burned gases, the early and
later images can easily be separated in most cases.

A calculated flame front is shown in white at the leading edge of the TDC
flame image. This line was derived by threshold processing the image; all pixels
below a chosen value are black and the rest are white. Laplacian edge detection
then provides a flame edge 1-2 pixels thick, from which obvious noise is removed
with the image processing editor. All flame fronts shown in this work have been
derived in this way.

ENGINE CONTEXT OF CYCLIC VARIATION

The data and discussion are organized in parallel with the stages of flame
development in an engine, described as combustion-progress regimes in the
previous work [9]. As combustion proceeds from ignition to extinction in each
cycle, effects that enhance or delay combustion combine into many different burn
histories. These effects can interact in ways that result in the same mass
burned fraction at one point in the cycle but quite different burn development,
leading to a great deal of statistical confusion when attempting to establish
cause and effect.

Two geometric effects dominate the behavior in this engine: the peripheral
spark plug, and the swirl that comes from the port configuration in use. The
peripheral spark implies a longer burn across the chamber than a central plug so
that for most of the cycle the flame geometry is determined by only the side of
the early flame kernel facing the center of the chamber. Also, the spark is more
sensitive to swirl near the wall where swirl velocities are higher. The low
swirl number for the configurations discussed in this work imply that the
in-cylinder flow will in general contain variable large scale motions, rather
than a consistent overall rotation of the flow and flame from swirl.

It is important to state that the results discussed are representative
single-cycle examples from a much larger set of operating data. Care has been
taken to avoid misrepresentation of the whole by the sample.

SPARK CYCLIC VARIATION

In most current spark-ignition automotive engines combustion is initiated by
a 1-2 as duration high-voltage inductive arc. Important characteristics of this
spark are its energy, location, and duration. The energy of the spark is not
important if that energy is well beyond a known threshhold value [13], except
possibly for very large energies. A millisecond spark duration has been found to
enable the spark plug to act as a source of continuous ignition of the mixture
passing by the plug at velocities higher than the initial flame velocity [14].
Spark location, both relative to the combustion chamber (location, protrusion,
and strap orientation) [15,16] and relative to the plug itself, especially with
respect to the ground strap, is important for the early flame kernel growth. The
ground strap is both an obstruction and a heat sink for the newborn flame.

Images taken with a slant view up into the head from a position opposite the
spark plug show the spark moving randomly around on the electrode (Figure 3a).
Similar motion of a pulsed spark occurs without flow and results in a
significant variation of the distance between the spark and the base of the
ground strap. The location of the arc on the plug structure and the shadowing of
the strap by the flame indicates that the reference image location is accurate
compared to the scale of the spark width.

Greater variations in spark location occur as a result of flow, as shown in
Figures b-e. In these figures, four sample spark images are each shown three
times relative to a reference box: by itself at the bottom left, magnified x2 at
the bottom right, and relative to the spark plug at the top. These spark loops
are extreme instances taken from a large set of spark images. The spark was
initiated at 40° ETD, and a 0.1° long exposure was taken
3.0° later. Looping of the extent shown in Figure 3 occurs for 1-5%
of the sparks, while perturbations on the order of the electrode width or larger
occur for approximately 20% of the sparks.

This looping of the spark is caused by flow through the gap that carries the
low inertia ionized gas column away from its original location. As the arc moves
and its channel length increases, the voltage needed to drive the arc increases
until it becomes large enough to break down a new arc across the original gap.
This characteristic voltage rise and sharp drop is a common characteristic of
inductive sparks in engines. Such a concept has been explored as a technique for
determining engine general flow parameters [17]. It should be noted that the
spark plug indicated in Figure lb does protrude vertically into the head volume,
where presumably higher flow velocities exist relative to a plug at the top
center of the chamber.

Figures 3b-d show loops from 1-port data in the direction of the mean swirl
for this case. Loops in the other direction do not occur in this configuration,
although the 1-port data with little or no mean swirl exhibit loops in both
directions. Figure 3e shows an example of a loop in the reverse direction from
the 1-port data. The actual frequency of occurrence of these loops in the engine
cannot be specified from the occurrence of loops in the images. Loop size and
direction will fluctuate in time due to the fluctuating flow; loop history
cannot be determined by the single view and single exposure time of this data.
However, the frequencies cited above are a lower limit. They are more accurate
in the Export case because the mean swirl lessens the dispersion in direction of
the loop.

The data in Figures 3b-d were taken at 1000 rpm for the 1-port case. Data at
500 rpm show similar trends, keeping the delay angle the same: 3o
after spark initiation. At 1000 rpm the time delay is half that at 500 rpm, but
velocities scale up by approximately a factor of 2 to give the same behavior in
the images. However, there should be twice as many breakdowns at the higher
speed and thus a higher effective ignition source size, especially if the flow
is constantly changing direction, as in the low-swirl case.

The existence of these loops in an engine implies a consequent variation in
the size of the ignition source from cycle to cycle, probably with a resulting
variation in initial flame kernel size. Since these loops are always associated
with relatively strong flows around the spark plug, only detailed measurements
can discover whether the flow or the larger spark are greater contributors to a
larger initial flame kernel growth. The data presented here do not have adequate
time development information to address the question.

The low frequency of occurrence of the loops does imply that there are many
cycles without loops, even compensating for possible loops that are developing
or that are out of the plane of view. Assuming that the loops begin to form when
the spark is initiated they imply a minimum flow velocity of approximately 5
m/s; a displacement of 3 mm over 3° crankangle at 1000 rpm. This
compares with a mean piston speed of 3 m/s and is consistent with in-cylinder
turbulence measurements in engines. Given a typical turbulence spectrum of from
400 to 1500 Hz, low velocity periods on the order of the spark duration are
frequent, supporting the cyclic variability in spark loops, and thus in
effective spark size.

This effective spark size will also vary with spark timing. The voltage
needed to break down the gap increases in proportion to the square root of the
pressure [17]. Thus the loop length attained before a new breakdown occurs
across the gap will increase significantly from 60° BTDC to 30°
BTDC, causing a larger cyclic variation in effective spark size for the later
spark. A smaller counteracting effect is that in-cylinder velocities decay with
time, leading to smaller loops later in the cycle.

EARLY FLAME KERNEL CYCLIC VARIATION

At the present time, changes in the early flame kernel, before there is
significant mass burned in the cycle, are thought to account for most cyclic
variation in spark-ignition engine combustion [3,16,18]. The two primary causes
of variation in this early kernel are thought to be distortion by the local flow
and interaction of the flame with the spark plug structure. Other factors in
cyclic variation include mixture nonuniformity (which will not be considered in
this work) and, for very lean operation, misfires.

The wide variety of flame behaviors that can occur during very lean operation
is illustrated in Figures Barb, which show a sequence of 9 consecutive cycles
from a 1-port firing run. For theme cycles the spark was initiated at 60°
BTDC, manifold absolute pressure was 0.75 atm. and the intake equivalence ratio,
øi, was 0.60. Slant side view exposures were taken at 57°
BTDC for a duration of 0.1° and at 20° BTDC for a duration
of 2.0°, while bottom view exposures were taken at 20°
BTDC for a duration of 2.0° and at TDC for a duration of 1.0°.
The images have been enhanced for maximum contrast and the background cylinder
patterns have been superimposed. Piston position at 20° BTDC is shown
in the side view. Cycle number in the sequence is shown at the upper left.
Bottom views were taken for each cycle but are presented only for the cycles
with combustion.

The images in Figures 4a-b show a great deal of dispersion in early flame
kernel size, shape, and location, but they show very little spark location
variation. Flame holding over an extended period is shown in the side view flame
image in cycle 5. Flame blowoff is shown by the side view in cycle 11, which is
an illustration of flame convection. Both effects have been seen in research
engines (e.g. [19]). The flow process of flame stretch is more difficult to
demonstrate unambiguously, but cycle 11 also seems to be an example of this,
based on the aspect ratio of the flame. The flame development in Cycle 7
indicates that flow velocities have been low throughout the growth of this early
kernel.

This set of sequential cycles clearly illustrates the wide variety of
mechanisms that give cyclic variation in the early flame kernel. The frequency,
or even the existence, of each effect changes in complicated ways with engine
configuration and operating condition. The critical parameters are the laminar
flame speed, gas expansion velocity, and local flow velocity.

The three-dimensional shape and location of the early kernel are all
important for later flame development. A compact flame is shown in cycle 7,
while an elongated flame is shown in cycle 11. Given that the early kernel grows
in all directions, the flame in cycle 11 will grow much faster than that of
cycle 7, resulting in a more rapid early burn phase. Such a correlation is
discussed below. Another way of describing this effect in current engine
terminology is to say that there is a shorter ignition delay time for the
elongated flame. Such terminology is misleading, since the actual phenomenon
involves early flame growth rate differences.

The location of the early kernel is important due to the delaying effect of
the spark plug structure. Figure 5 shows the variety of positions taken by the
early flame kernel. Figures 5a-b and 5c show the flame behind and in front of
the plug, as indicated by the shadow or lack of it in the flame image. Figure 5d
shows the flame (and spark) blown to the right of the spark plug. In the 1-port
flow in this engine configuration the flame casting the shadow is near the wall
and therefore must pass the spark plug again before it fills the chamber. This
effect is more important if the clearance space is small or the kernel is still
small when it reaches the plug again.

For the spark advances used in this work, the clearance volume is large at
the time of early kernel growth, so the location of the early kernel with
respect to the spark plug is less important than changes in the kernel size and
shape. However, for the previous work [9], when the spark is fired at 5°
BTDC the clearance between the piston and spark plug is much smaller, the effect
should be evident. This seems to be the cause of the dent in the kernel near the
spark plug of some cycles shown in that work.

These flow effects all occur for the 1-port case with a mean swirl ratio of
1. The mean rotation of this flow in any specific cycle, however, varies from
approximately zero to more than 1, implying that the mean swirl ratio for such a
low swirl number is a misleading indicator of the flow in that particular cycle.

The true misfires shown by the even numbered cycles in Figures 4a-b are
complementary data to that discussed in the previous work [9], showing that for
late ignition flames exist even though there is no measurable pressure rise from
combustion. For similar very lean mixtures, compression of the gas by the piston
can make that mixture ignitable. The extreme spark advance of the data in
Figures 4a-b begins ignition at much lower pressures and lower temperatures
compared with conditions that would occur near TDC.

As an extreme form of cyclic variation the sequence shown in Figure 4 is a
sample of an oscillation between true misfires (no flame) and complete burns.
This is another case of the residual gas oscillation shown and discussed in the
previous work [9]. Air-fuel mixture in the residual gas following a misfire
cycle makes the overall mixture rich enough to burn completely in the next
cycle, while the situation is reversed following a burned cycle. Such firing
conditions can only be maintained in a motored engine because cyclic pressure
variation is so severe.

This is a prior cycle effect that has been seen extensively in lean operation
in this engine. Partial burns result in stronger burns in the following cycle in
general. In real engines such partial burns can occur at any average
stoichiometry for a variety of reasons, ranging from spark plug fouling,
non-uniform air-fuel charge, to any flame that is slow enough to be quenched by
volume expansion arising from piston motion or exhaust valve opening.

MID-CYCLE BURN CYCLIC VARIATION

Flame images taken through the piston window of 24 sequential cycles from a
1-port MBT (Minimum spark advance for Best Torque) run at 500 rpm are shown in
Figures 6a-b. The spark was initiated at 40° BTDC and the
superimposed exposures were taken at 20° BTDC for a duration of 2.0°
and at TDC for a duration of 1.0°. The early exposure is shown as the
bright part of the image around the spark plug. The engine was operated at an
intake equivalence ratio of 0.75, and a MAP of 0.75 atm. MBT timing was
indicated by a Mass Burned Fraction (MBF) of approximately 60% at 10°
ATDC. As with production engine operation, there is considerable variation
around this mean. Simultaneous perpendicular side-view images were also taken
through the liner but these were views of the narrow piston clearance volume
without a view into the head, and are not presented here.

It should be reiterated that these images represent line-of-sight integrals
of the radiation from the flame. Thick flame fronts in the images do not
necessarily imply a thick reaction zone. However, at TDC the height of the
chamber is much less than its width, so that the major variations in the flame
shape at TDC are radial rather than axial. This is not true of the early flame
kernel, as discussed previously.

The flame images taken at TDC in Figure 6 show both a broad dispersion in
flame shapes-and flame progress. These images show flame shapes that are very
similar to those reported in the previous work [9], which were taken at similar
conditions except for a very retarded spark. The large-scale shape of the flame
fronts in these cycles range from spherical (cycle 1) to flat (cycle 22), to
severely indented (cycle 13). The flames tend to be rotated clockwise relative
to the location of the spark plug, consistent with the swirl ratio of 1.

The flame shape at TDC is similar to the shape of the side of the early flame
kernel facing the main part of the chamber in nearly all of the bottom view
flame images taken in the engine. This is a result of the dominance of the gas
expansion velocity as a part of the flame propagation velocity. Cyclic
variations in the early kernel geometry can thus affect the burn throughout the
cycle through the shape similarity, since different large scale flame geometries
will result in major variations in overall combustion, as shown below.

Figures 7a-b show 24 sequential cycles of 2-port flame images for conditions
identical in other respects to those of Figure 6a-b. Flame shapes are very
different, without any flat or indented flames. Combustion progress is slower in
general, as indicated by the flame areas, presumably due to a lower turbulent
flame velocity from lower turbulence levels. No swirl is indicated, as expected.
Once again there is shape similarity between the early and later flames so that
the flame shape throughout most of the cycle is determined by the effects that
caused the shape of the early flame kernel.

Given the prominence of this apparent flow effect on the flame, engine
operation was changed to find out how the flame shapes were affected. In Figures
8a-b flame shapes are shown for an increased intake equivalence ratio of 0.90.
Spark timing is 30° BTDC and the exposures occur at 20o
BTDC and TDC, as before.

As expected, these images show mostly round flame fronts, because both the
laminar flame speed has significantly increased relative to the in-cylinder flow
velocities. Camera saturation leading to vertical blooming causes the bright
distortions such as that seen at the lower right in cycle 22. There is evidence
of the presence of the secondary flow at a reduced level, seen in cycles 14, 22,
and 34. The reduced effect of the flow is well known statistically [7]; cyclic
variation is reduced as the mixture more closely approaches stoichiometric.

Figure 9 shows flame images from consecutive cycles taken at identical engine
conditions as the data of Figures 8a-b, except that the exposures were delayed
10° to be at 10° BTDC and 10° ATDC. For cycles
10 and 16 the early and later images are shown separately, since they happened
to fall on either side of a video frame time boundary. Although the flames have
often progressed beyond the view of the piston window, cycles 8, 11, 12, and 15
definitely show the indentation characteristic of the secondary flow so
prominent in leaner mixtures. This shows that the flow effect is still present,
but it occurs later in the cycle and has a reduced effect on the flame shape.
That the flow field in the unburned fluid is compressed in front of the flame by
burned gas expansion behind the flame has been shown (Witze [20]). These flows
can affect flame shape and mass burned in the cycle even at near-stoichiometric
mixtures because while the effect occurs late in the volume burned history, it
occurs for the middle and late stages of the mass burned history.

As an incidental point of interest, the bright spots in the images are caused
by some unknown particle or drop inside the volume of the chamber, usually
behind the flame front. Because the engine uses gaseous fuel and has no oil
behind its compression rings, there is no obvious explanation for their presence
in the images. There is also no evidence of oil leaking through the valve
guides.

MASS BURNED CORRELATIONS - Such a large dispersion in flame shape as seen in
Figures 6a-b would be expected to cause related changes in MBF history. Flame
areas derived from flame images such as those presented here imply mass burned
rates, and individual cycle MBF's have long been analyzed in this way [6,11].
Although unique flame shapes such as that of cycle 13 might be expected to
result in a unique MBF shape, this was not the case. Initial attempts at
correlating flame shape and MBF failed.

To discover more subtle correlations, MBF curves were classified relative to
an Beverages cycle, taken to be cycle 9, shown in Figure 2. Combustion progress
was catalogued as having early, average, or late timing relative to the average
cycle at 1%, 50%, and 90% MBF. The data of the cycles shown in Figures 6a-b were
then examined to discover any relationships between flame size or shape, and
mass burned progress.

One important correlation that could be identified was that cycles whose
early flame kernels were larger correspond to those cycles with an early rise in
mass burned. Cycles 7,10,13,14,16,17 and 19 have large early kernels as seen by
the bright region around the spark plug. Compare, for instance, the early flame
kernels of cycle 6 and cycle 7. The sizes implied in Figures 6a-b can be
ambiguous due to the restricted view through the piston window, but the
simultaneous side view through the sapphire was used to confirm the extent of
these early flame kernels, while eliminating Cycle 8 because it did not extend
to the cylinder wall. The correlation was 1 to 1: early rise in MBF occurs if
and only if there was a large early flame kernel. MBF curves for two of these
cycles are shown compared with the average cycle in Figure 10.

This correlation is necessarily true at the instant that the flame image is
recorded, because the volume of the kernel is a measure of the mass burned at
that time. However, while these images were taken at negligible mass burned
fraction, they correlate with a larger mass burned rate throughout the early
stage of combustion after the images were taken. This implies that whatever
subsequent combustion effects exist, they are of second order compared to the
effect (presumably the local flow) that causes the dispersion in the very early
flame kernel size.

Another major correlation that became apparent was that cycles with flat
flame fronts across the cylinder have lower mass burning rates than those with
round flame fronts. This is seen in the MBF curve shape as a decreased slope and
continually delayed timing with respect to the average. Cycles
3,4,7,10,14,19,22,and 23 are such cycles, once again using the side view to sort
out potential candidates. This correlation was also 1 to 1: the mass burned rate
is slower throughout the cycle if and only if the flame front is flat. MBF
curves for sample cycles relative to the average cycle are shown in Figure 11.

These two effects combine when present together to confuse any global
statistics. Some of the large early kernels result in flat later flames. These
flames are early relative to the average cycle at the start of significant mass
burned, and late at the end of the mass burned, as shown by Cycle 7 in Figure
10. The unusual shape of Cycle 13 is seen to be early throughout the cycle
relative to the average cycle, presumably because the flow distortion has caused
the flame to grow in two separate large sections, resulting in a larger total
burning rate.

Combining these effects with the fact that the shape of the early flame kernel
usually determines the shape of the mid-cycle flame, it can be concluded that
the shape of the side of the kernel facing the main part of the chamber will
determine whether the flame develops as a flat or round flame. This can be seen
to be mostly true by examining the small and bright early flame kernels from the
first exposure of the images of Figure 6a-b. That this is not uniformly true is
due to the influence of strong secondary flows of the scale of the chamber as
shown in cycle 13, and discussed below.

FLAME FRONT ANALYSIS - To further explore the flame shapes in the flame
images, flame front locations were derived by assuming the flame to begin at a
fixed intensity level in the images. Since the flame edges are usually sharp, a
change in threshhold level moves the front slightly, but does not change its
basic shape. These are not actual flame front loci, but the sum of all of the
axially varying protrusions that occur. This would lead to ambiguity for any
detained conclusions, but probably does not change the perception of overall
shape, because the flames are basically axial. The spark is not high above the
piston, resulting in flames that reach the piston early in their history and
then quickly lose their axial curvature.

The illustrations shown in this paper are significantly degraded from those
used in data analysis. The original flame images used to derive the flame fronts
are of clearly better quality. For a few of the flame fronts the original image
was used to complete the locus of the front, where the background light
prevented accurate threshholding in part of the image. The flame fronts shown in
Figures 13-18 are presented using three different line widths to show the flame
fronts for up to six cycles, leading to some confusion of the individual flame
fronts. The original analysis was done by assigning each flame front a primary
color; each one of the superimposed six flame fronts was easily distinguished
and could be manipulated separately using its color index.

For the 1-port case, flame fronts were derived from the images shown in
Figures 6a-b, and are given in Figure 12 together with the outlines of the spark
plug ground strap, the cylinder bore, and the piston window border. The cylinder
pattern location is accurate to approximately the scale of the line-widths used.

The 24 cycles of Figure 12 show a band of flame front locations that might
represent the cyclic variation of combustion at these conditions. However, since
the MBF arises from the volume enclosed by the flame, very different shapes
imply the same mass burned. Note that the MBF at TDC is 25-30% (Figures 10 and
11) whereas the volume burned is 50-60%.

The effect of the mean swirl can be removed to better compare the flame
shapes in different cycles. This image analysis assumes solid body rotation
about the axis of the cylinder. A line is drawn through the flame front so that
the area between the flame front and the line on either side of the line are
estimated to be equal. The slope of this line relative to a line bisecting the
cylinder and passing through the spark electrode was used to give a rotation
angle of the flame relative to a spherical expansion from the spark electrode.
The flame fronts were then rotated back through this angle so that each one
appears to have expanded from the spark plug.

The apparent dispersion of flame front location is significantly reduced by
this processing, and these images are a better representation of the real MBF
dispersion. In order to examine the question of the relative speed of round vs.
flat flames, these flame fronts were sorted by shape. The sorting was done on
the qualitative basis of how well their shape approximated a circular arc with
its center at the spark plug, or whether the shape showed the characteristic
central indentation toward the spark plug. Cycles that did not fit well in
either category were classified as "intermediate".

Figures 13a-d shows the result of this sorting. Figures 13a shows a set of
round flames with very small location dispersion, while Figure 13b shows
indented shapes, also with very small dispersion. The indented shape became
apparent when the flatter flames were grouped together. Figures 13c-d show the
remainder of the flame fronts, whose shape seemed to be intermediate between the
round and indented shape.

The groups form three nearly equal sets, and the very small dispersion of the
round and indented sets very strongly implies the existence of a bistability in
shape. This is presumably due to an unstable large-scale secondary flow in the
cylinder. The indented flames indicate its presence, while the round flames
imply flame development when this flow is not in the cylinder. The effect of the
flow varies in strength, but is most prominent in cycle 13. This phenomenon was
previously seen for similar conditions with late firing in the previous work
[a].

From this grouping of cycles, the relative burn rates of round vs. flat
flames can be assessed quantitatively, as shown in Figure 14. Peak burn rate was
derived by differentiating MBF data, derived from pressure data, and plotting
the maximum. The peak burn of the cycle was chosen rather than the fraction
burned at a particular time to lessen the influence of cycle phasing on the burn
rate correlation with the different flame shapes.

The results in Figure 14 show that round flames have significantly higher
peak burn rates than flatter flames, The question mark for cycle 21 indicates
that the flame shape at TDC could not be determined because the flame front
extended beyond the view of the piston window.

However, it is almost certainly round, as indicated by the shape of the early
flame kernel, consistent with the burn rate correlation. Further confirmation of
the correlation is that cycles 3 and 13 are extremely dented, to form two
separate essentially round flames so that the cycle as a whole has a higher burn
rate.

The cause for the higher burn rate of the round flames is that for a flame
propagating normal to its surface, a circular flame increases its area with
time, whereas the area for the flat flames remains constant. This is a simple
geometric argument supported by the data of Figure 14. The implication of this
fact is that the smaller scale wrinkling of the turbulent flame, while it is
certainly present, is relatively constant from cycle to cycle. A major cause of
cyclic variability for this engine configuration at these operating conditions
is an unstable large-scale secondary flow.

An alternate possibility as a cause for such distorted flames might be
non-uniform mixing of residual gases. The intake fuel and air in this experiment
are extremely well mixed, supplied to the intake tank by sonic orifices through
an L/D > 100 tube. To eliminate the possibility of residual effects, the
engine was fired on alternate cycles at a reduced intake equivalence ratio to
give the same pressure release. No change was seen in the flame shapes.

To learn whether the flame shapes (and thus the secondary flow) are somehow
connected with the swirl in each cycle, the swirl angles used to derive the
corrected flame fronts are plotted in Figure 15. No correlation emerges. The
data are consistent with the average swirl ratio of 1, but there are three
cycles without any indication of swirl. Even though there is mean directed flow
from the placement of the intake duct, this does not necessarily result in
swirl, another indication that mean swirl numbers up to 1 are a misleading
descriptor of the detailed flow.

Engine speed was another change made in operating condition to explore flame
shapes. Flame images were taken at 1000 rpm at the same conditions as those for
the flame images of Figures Barb. Flame front locations were derived as before,
and then corrected for swirl, as before, with results presented in Figure 16.

As can be seen in the flame front data, the bearable flame shapes have almost
entirely disappeared. Replacing the two shapes is an intermediate shape, so that
the average remains the same while the dispersion has decreased. These flow and
flame changes with rpm would thus not appear in a global statistical analysis of
the cycles based on heat release.

A fundamentally different flow in the engine occurs during Export operation.
The pressure drop across the valves is much lower, as are the intake velocities,
and there is also no significant swirl.

Figure 17 shows the flame fronts derived from the images of the 2-port, øi
= 0.75, 500 rpm, case shown in Figures 7a-b, and these possess significant
dispersion. The shapes of these fronts suggest another sorting such as was done
above, this time into elongated and round flame fronts as shown in the examples
in Figures Arab. The sorting is not so effective as in the previous case, but it
does seem to again imply a bistability between round flames and elongated
flames. It should be noted that the elongation is aligned with the pentroof, and
is consistent with preferential expansion of the flame along the direction of
greatest area. The round flames seem to have some clockwise swirl, probably
preventing the alignment of the flame expansion with the pentroof axis. The
importance of the effect of large-scale unstable secondary flows seems to be
reinforced by this data.

The flame variation that is characteristic of Export operation at 500 rpm in
this engine can be seen in the bottom view images. These images show flame front
variation under leaner conditions than discussed above, and support the
flow-flame interaction conclusions already drawn from the data.

The side view images of combustion in each cycle show the location of
continuing combustion at the wall. Fully burned regions result in uniform
intensity, while radiation from the burning turbulent flame gives regions of
highly structured intensity. The faint striations aligned with the piston-top
line are caused by compression ring debris on the inner wall.

The shape and progress of the flame shown in the bottom view at TDC, results
in large differences in the size of the regions of continuing combustion along
the wall 50° later in the cycle. Since the spark is opposite the view
of the camera taking the side view images, these images show the burning flame
area as the flame approaches the wall. Combustion is complete at 50o
ATDC in cycles 10 and 17; in cycle 10, this is due to the relatively greater
progress of burning, compared to other cycles, but in cycle 17 combustion is
completed because the flame shape happens to closely conform to the shape of the
chamber when it reaches the wall. The flatter flames of cycle 13, 19, 24, and 25
cause longer combustion durations compared to the rounder flames due to the
flame-chamber shape mismatch, even though the flame areas are comparable at TDC.
This effect adds to the slower burning rate of the flatter flames to give even
larger cyclic dispersion between flat and round flames.

The degree to which the flame shape conforms to the combustion chamber shape
is another major cause of cyclic variation. It is one that should be nearly as
important for central ignition as for peripheral ignition due to flow distortion
of the early kernel shape and the constancy of that shape.

Another cause for cyclic variation can be identified as axial variations of
flame shape. Almost all of the side-view images indicate that the flame front is
vertical, supporting the earlier assumption that the bottom view images are good
representations of flame shape even though they are axial averages. However,
cycle 23 shows significant slope to the flame front and thus a longer time to
complete combustion. As with all of the other effects causing cyclic variation
this will increase the statistical dispersion of the cycles, although in any
particular cycle it may not cause delayed combustion compared with the average
cycle, depending on prior events.

It is interesting to note the variations of the very early flame kernel that
are shown in the side view in the combustion completion images. While most of
the kernels are displaced to the right, in the direction of the mean swirl, the
flame images for cycle 25 show the kernel developing between the spark plug and
the wall. This is consistent with the above results concerning the dispersion in
the shape and location of the early flame kernel. Cycle 25 presumably is a case
where the early kernel is blown toward the wall, and is stretched out because of
flow divergence there, leading to the flat flame that later develops.

A striking aspect of the side view images is that the intensity structure
scale is very uniform from cycle to cycle. This is surprising because there is
such a large variation in overall shape and large scale flow between cycles.
Apparently the turbulence at these scales varies little, which is consistent
with the correlation discussed above that indicates that large scale changes in
flame shape account for the primary changes in overall flame area for most of
the cycle. Similar length scales have been found with particle imaging
velocimetry in an engine by Reuss et. al. [21].

COMPARISON WITH PREVIOUS RESEARCH

It is appropriate to briefly discuss how the results presented in this work
interleave with some relevant previous research.

The recurring theme (e.g.[22]) of the usefulness of decreasing overall
combustion time to reduce cyclic variability has interfered with the
identification of the fundamental causes of cyclic variation. Massive regression
analyses have been performed on engine operating statistics to relate engine
parameters to cyclic variation [7]. Such studies show that absolute cyclic
variability can be lessened by finding the means to concentrate pressure release
more near TDC. However, the causes of cyclic variation result in a relative
dispersion of combustion duration compared with the mean, and this relative
dispersion may increase or decrease as the combustion duration is shortened. An
example is to assess a parameter such as equivalence ratio, which has its major
effect on cyclic variability by accelerating combustion in general, primarily
through an increase in laminar flame speed [23,24]. However, the higher
equivalence ratio may be concentrating the cyclic variation at the end of the
cycle where engine performance can be degraded by knock.

Another significant part of previous work on cyclic variability depends on
the use of ionization gap measurement of flame progress [1]. These measurements
are only reliable for spherical flames, and this work shows that such flames do
not occur in general for lean mixtures or beyond the midpoint of volume
combustion. For example, in two consecutive cycles with a flat, then a round
flame, an ionization gap measurement taken opposite a peripheral plug ignition
would indicate a spuriously low flame speed for the flat flame. Ionization gap
work may often be misleading based on 3-dimensional flame shape variations, and
must be evaluated carefully.

Cyclic variation was originally associated with the variability of cycle
pressure release (e.g. [25]). The approach of studying combustion progress by
analyzing cycle pressure history has been deliberately avoided in this work. The
causes of cyclic variation have been found to be such that each one gives a
separate dispersion in combustion progress, combining to result in a dispersion
in combustion phasing whose cause will often be nearly impossible to unfold from
statistical analyses of cycle pressure measurements.

Flow in the cylinder is generally accepted to be the primary cause of cyclic
variation in engines, but only by inference, due to a lack of measurements [1].
For the problem of in-cylinder flow, the boundaries are predetermined, if not
fixed, and the decaying turbulence has a particular kind of initiation in terms
of energy, duration, and geometry. This leads to repeatable characteristics of
the flow, such as mean swirl and turbulence level. The primary effect of the
flow is thought to be through the variation in the early flame kernel (e.g.
[2,3,16,181). This is confirmed by the present work, identifying the cause
primarily as the determination of the size and shape of the early kernel, which
usually determines the shape of the later flame and consequent geometric effects
on burn rate.

A number of effects have lead to confusion in the isolation of the causes of
cyclic variability in the early flame kernel and later in the cycle. The work of
zur Love and Braces [26] has revealed some new characteristics of engine
turbulent combustion, but since each image was taken in a different cycle,
limited information can be deduced about flame development in each cycle or
cyclic variability. The work of Keck et. al. [16], while clearly showing the
variation in 3-dimensional early flame kernel size, shape, and location similar
to that shown here, employs an square engine geometry (and lower compression
ratio) that must change the flow fields in major ways. A spherical flame
expansion model is forced onto a flame that was shown have large changes in
aspect ratio. While flame shape similarity between the early and later flames
can be seen in many flame images in the literature [e.g. 33] it is often masked
by 3-dimensional effects and when the flame is nearly spherical.

Current experimental work [27,28] has also identified secondary flows called
tumble, that seem to make predictable contributions to turbulence
characteristics at TH. Such flows are highly suggestive of the bearable,
large-scale flows discussed in this work. Both of these effects are seen in a
similar form and used as simple combustion quality diagnostics using flame
imaging in a manner much like the present work by Nakamura et. al. [29]. This
and other evidence indicate that small scale turbulence need not be the major
contributor to cyclic variation, as has been claimed [5].

Engine speed variations have been found to have a mixture of effects on
cyclic variability [1], although increased turbulence with speed would be
expected to increase cyclic variability through greater variation in the flow
when the early flame kernel is growing. The disappearance of unstable secondary
flows at higher speeds may be one effect that causes the cyclic variability to
decrease with speed.

The effect of residual gases on cyclic variability has been studied by a
number of authors [30,31,32]. Statistical studies of sets of engine cycles do
indicate some prior-cycle dependence at very lean conditions [30]. These cycle
conditions involve slow flames and long combustion times compared with
stoichiometric conditions. These are conditions where one would expect a
significant number of partial burns, lending support to the effect of partial
burns seen so often in lean combustion in this engine. A combustion instability
arising from the interaction of laminar flame speed and combustion duration with
exhaust temperature and residual fraction has not been observed. Such an effect
has been proposed [31,32], but is at least obscured by other more obvious
effects.

The need for the flame to conform to the combustion chamber is well known
(e.g. [33,34]); it is a basic rule in current engine design. The interaction of
variable flame location and shape with chamber shape is also known (e.g.
[2,29]), but since there is currently no way to control flame positioning,
little work has been done on this aspect of cyclic variability.

CONCLUSIONS

The data presented and discussed above lead to a variety of conclusions about
the nature and causes of cycle-by-cycle variation in SI four stroke engines.

1) Demonstrated causes of cyclic combustion variation:

i) A faster burn rate throughout the early stages of combustion correlates
with the size of the very early flame kernel; earlier rise of the MBF occurs if
and only if there is a large very early flame kernel. These early flame kernels
are larger at negligible MBF, indicating that the important mechanism that
causes the larger relative size occurs very soon after flame initiation and is
not overcome by other later effects in the first part of the burn.

ii) The early flame kernel is usually shaped by the local flow, causing large
variations in kernel size, shape, and location, and consequent variation in
combustion through the control of the later flame geometry by the geometry of
the early kernel.

iii) Flame shape is usually self-similar as the flame grows after the early
flame kernel stage, modified by wall-interaction effects.

iv) Flat flame fronts have a lower mass burned rate than circular flame
fronts, and cause relative combustion progress delays. These flat flames arise
both from distortion of the early flame kernel and from interaction of the flame
with large scale flows. The change between flatter and rounder flames is a major
cause of overall cyclic variation of combustion

v) Changes in the dares to which the mid-cycle flame shape conforms to the
combustion chamber shape is another major cause of cyclic variation. Major
differences in the distance between the flame front and the wall around the
circumference of the flame cause a dispersion in a combustion duration due to
the differences in the necessary flame travel time before combustion completion.
This is a factor that should be nearly as important for central ignition as for
peripheral ignition, due to flow distortion of the early kernel shape and the
constancy of that shape.

vi) Another cause for cyclic variation can be identified as axial variations
of flame shape. This is a comparatively rare event that also causes combustion
delay because the flame shape does not conform to the combustion chamber
axially.

vii) For Export operation in this engine at 500 rpm, major cyclic dispersion
is characterized by flame front shapes that arise from 2 basic shapes that
themselves have much less dispersion: round and indented. The existence of this
separation strongly implies a bearable largescale flow in the cylinder.

viii) The lean combustion oscillation found in earlier experiments [9] also
occurs at advanced spark timing; cycles alternate between complete combustion
and total misfires.

ix) Partial burns that cause stronger burns in the following cycle are
significant causes of cyclic variation in lean combustion.

i) The existence of looping of the spark away from the plug gap causes an
apparent cyclic variation in the ignition source size. The spark loops are
caused by local flow through the spark plug gap and can be used to estimate that
flow velocity.

3) Conditions demonstrated to affect cyclic variation of combustion:

i) The unstable large-scale secondary flows disappear at 1000 rpm, where the
average flame shape becomes that seen in each cycle. This is important evidence
of a fundamental change in flow and combustion behavior with engine speed.

ii) For equivalence ratios approaching stoichiometric the large scale
secondary flows affect flame development much less. However, these flows do
become important again late in the combustion event.

iii) The cyclic variations of combustion caused by the interaction between
the flow and the flame are magnified by peripheral ignition. This is due to the
long flame travel required to complete combustion, one-sided flames, and the
early interaction with the chamber wall.

4) Miscellaneous conclusions:

i) Relative mass burning rates between cycles correlate quantitatively with
whether the large scale flame shape is round or flat.

ii) Low swirl numbers may be a misleading descriptor of the flow, since the
mean rotation of this flow in any particular cycle varies from approximately
nothing to more than that indicated by the swirl ratio.

iii) In this engine configuration there is significantly less variation in
flame shape and flame speed during 2-port operation compared with 1-port
operation.

iv) The spark plug functions as a flame holder at low swirl ratios, except
for very lean mixtures when blown-off flames sometimes do occur.

ACKNOWLEDGEMENTS

Numerous discussions with Rod Bask on the various implications of the flame
image data were very helpful. Group meetings with interested staff members of
the Fluid Mechanics Department and the Engine Research Department also provided
useful criticism and discussion.